Power Tool

ABSTRACT

[Problem] To provide an effective technique for further improving the vibration reducing performance of a power tool. 
 
[Means for Solving the Problem] The invention provides a power tool which includes a tool bit, a first operating mechanism that linearly drives the tool bit and thereby causes the tool bit to perform a predetermined operation, and a dynamic vibration reducer  151  that reduces vibration in the operation of the tool bit. The dynamic vibration reducer includes a weight  153  that can linearly move under the action of biasing forces of a plurality of biasing springs  157  acting upon the weight  153  toward each other. The weight  153  is driven by a second operating mechanism  116  that mechanically excites at least one of the elastic elements, via the elastic element.

FIELD OF THE INVENTION

The present invention relates to a technique for reducing vibration in apower tool, such as a hammer and a hammer drill, which linearly drives atool bit.

BACKGROUND OF THE INVENTION

Japanese non-examined laid-open Patent Publication No. 52-109673discloses an electric hammer having a vibration reducing device. In theknown electric hammer, a vibration proof chamber is integrally formedwith a body housing (and a motor housing) in a region on the lower sideof the body housing and forward of the motor housing. A dynamicvibration reducer is disposed within the vibration proof chamber. It isdesigned such that the dynamic vibration reducer reduces a strongvibration which may be caused in the longitudinal direction of thehammer when the hammer is driven

In the above-described dynamic vibration reducer, the weight is disposedunder an action of the biasing force of an elastic element. The dynamicvibration reducer performs a vibration reducing function by the weightbeing driven according to the magnitude of vibration inputted to thedynamic vibration reducer. Specifically, the dynamic vibration reducerhas a passive property that the amount of vibration reduction by thedynamic vibration reducer depends on the amount of vibration. In actualoperation, a considerable load is applied from the workpiece side to thetool bit, for example, by a user performing the operation while stronglypressing the power tool against the workpiece and therefore, vibrationreduction is highly required in such case. However, in some cases, theamount of vibration to be inputted to the dynamic vibration reducer maybe reduced.

SUMMARY OF THE INVENTION

Object of the Invention

It is, accordingly, an object of the present invention to provide aneffective technique for further improving the vibration reducingperformance of a power tool.

Resolution to Solve the Problem

The above-described problem can be solved by the features of claimedinvention.

The present invention provides a power tool which includes a tool bit, afirst operating mechanism that linearly drives the tool bit and therebycauses the tool bit to perform a predetermined operation, a dynamicvibration reducer that reduces vibration in the operation of the toolbit via a weight that reciprocates under the action of a biasing forceof an elastic element, and a second operating mechanism thatmechanically excites the elastic element to thereby forcibly drive theweight.

The tool bit typically comprises a hammer bit that performs a hammeringoperation or a hammer drill operation on a workpiece, or a saw bladethat performs a cutting operation on a workpiece.

In the present invention, in a passive vibration reducing mechanism inthe form of the dynamic vibration reducer, the weight is actively drivenby the second operating mechanism. Therefore, regardless of themagnitude of vibration that acts on the power tool, the dynamicvibration reducer can be steadily operated. As a result, a power tool isprovided which can ensure a sufficient vibration reducing function evenin such operating conditions in which only a small amount of vibrationis inputted to the dynamic vibration reducer and the dynamic vibrationreducer does not sufficiently function.

According to this invention, particularly, the second operatingmechanism is provided for mechanically exciting the elastic element thatapplies a biasing force to the weight. Thus, the timing of excitationcan be appropriately adjusted, and the phase of linear motion of theweight can be freely set. Therefore, the timing for driving the weightcan be caused to coincide with the time at which an impact force isgenerated during operation of the tool bit, so that vibration reductionby the dynamic vibration reducer can be optimized.

In the dynamic vibration reducer, either one or more elastic elementsmay apply a spring force to the weight. In the latter case, at least oneelastic element needs to be mechanically excited.

The dynamic vibration reducer may have specific damping characteristicssuch that the behavior of the dynamic vibration reducer is stabilized.Specifically, the dynamic vibration reducer may be provided with suchdamping characteristics that the amplitude of the weight varies within aspecified amplitude range in a predetermined region of frequencies ofexcitation by the second operating mechanism and that the phasedifference between the weight and the second operating mechanism varieswithin a specified phase difference range in the predetermined frequencyregion. In other words, it is preferable that the degrees of variationsin the amplitude of the weight and the phase difference fall withinrespective specified ranges in a predetermined region of frequencies ofexcitation by the second operating mechanism. Vibration reduction by thedynamic vibration reducer becomes effective when the predeterminedregion of excitation frequencies covers the actual operating frequencyregion which is set allowing for variations in manufacturing or in useof the reciprocating power tool. Such construction is particularlyeffective in appropriately alleviating errors in manufacturing or inuse, such as variations in the elastic coefficient of the elasticelement of the dynamic vibration reducer, an error in the mass of theweight, and variations in operating frequency.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a sectional side view schematically showing an entire electrichammer drill according to a first embodiment of the invention.

FIG. 2 is a sectional plan view showing a dynamic vibration reducer anda second motion converting mechanism that forcibly excites the dynamicvibration reducer, with a biasing spring under pressure.

FIG. 3 is a sectional plan view showing the dynamic vibration reducerand the second motion converting mechanism that forcibly excites thedynamic vibration reducer, with the biasing spring under no pressure.

FIG. 4 illustrates a forcible excitation model with a damping element.

FIG. 5 is a graph of an “embodiment” showing the relationship between acoefficient ρ (−) of the amplitude of a weight 153 and an excitationfrequency f (Hz) and the relationship between a phase difference θ (°)between the weight 153 and the vibration input and the excitationfrequency f (Hz).

FIG. 6 is a graph of a “comparative example” with respect to the“embodiment” shown in FIG. 5.

FIG. 7 is a sectional side view schematically showing an entire electrichammer drill according to a second embodiment of the invention.

FIG. 8 is a sectional plan view showing a dynamic vibration reduceraccording to the second embodiment and a second motion convertingmechanism that forcibly excites the dynamic vibration reducer, with abiasing spring under maximum pressure.

FIG. 9 is a sectional plan view showing the dynamic vibration reduceraccording to the second embodiment and the second motion convertingmechanism that forcibly excites the dynamic vibration reducer, with thebiasing spring under medium pressure.

FIG. 10 is a sectional plan view showing the dynamic vibration reduceraccording to the second embodiment and the second motion convertingmechanism that forcibly excites the dynamic vibration reducer, with thebiasing spring under no pressure.

FIG. 11 is a sectional view taken along line IIX-IIX in FIG. 9.

FIG. 12 is a sectional view taken along line IX-IX in FIG. 9.

FIG. 13 is a schematic view showing the construction of a modificationto the first embodiment.

FIG. 14 is a schematic view showing the construction of a modificationto the second embodiment.

REPRESENTATIVE EMBODIMENT OF THE INVENTION First Embodiment

A first embodiment of the present invention will now be described withreference to FIGS. 1 to 3. In this embodiment, an electric hammer drillwill be explained as a representative example of a power tool accordingto the present invention. FIG. 1 is a sectional side view schematicallyshowing an entire electric hammer drill 101 as a first embodiment of thepower tool according to the present invention. As shown in FIG. 1, thehammer drill 101 of this embodiment mainly includes a body 103 and ahammer bit 119 detachably coupled to the tip end region of the body 103via a tool holder 137. The hammer bit 119 is a feature that correspondsto the “tool bit” according to the present invention.

The body 103 includes a motor housing 105, a gear housing 107, a barrelsection 117 and a handgrip 109. The motor housing 105 houses a drivingmotor 111 and the gear housing 107 houses a first motion convertingmechanism 113, a power transmitting mechanism 114 and a second motionconverting mechanism 116. The barrel section 117 houses a strikingelement 115 that includes a striker 143 and an impact bolt 145. Therotating output of the driving motor 111 is appropriately converted intolinear motion via the first motion converting mechanism 113 andtransmitted to the striking element 115. Then, an impact force isgenerated in the axial direction of the hammer bit 119 via the strikingelement 115. Further, the speed of the rotating output of the drivingmotor 111 is appropriately reduced by the power transmitting mechanism114 and then transmitted to the hammer bit 119. As a result, the hammerbit 119 is caused to rotate in the circumferential direction. Further,the rotating output of the driving motor 111 is appropriately convertedinto linear motion via the second motion converting mechanism 113 andinputted into a dynamic vibration reducer 151, which will be describedbelow, as a driving force for forcibly exciting the dynamic vibrationreducer 151. The first and second motion converting mechanisms 113 and116 are features that correspond to the “first operating mechanism” andthe “second operating mechanism”, respectively, according to thisinvention.

The first motion converting mechanism 113 includes a driving gear 121that is rotated in a horizontal plane by the driving motor 111, a drivengear 123 that engages with the driving gear 121, a crank plate 125 thatrotates together with the driven gear 123 in a horizontal plane, a crankarm 127 that is loosely connected at one end to the crank plate 125 viaan eccentric shaft 126 in a position displaced a predetermined distancefrom the center of rotation of the crank plate 125, and a drivingelement in the form of a piston 129 mounted to the other end of thecrank arm 127 via a connecting shaft 128. The crank plate 122, the crankarm 127 and the piston 129 form a crank mechanism which is a featurethat corresponds to the “first driving mechanism” according to thisinvention.

The power transmitting mechanism 117 includes a driving gear 121 that isdriven by the driving motor 111, a transmission gear 131 that engageswith the driving gear 121, a transmission shaft 133 that is disposedcoaxially with the transmission gear 131 and caused to rotate in ahorizontal plane together with the transmission gear 131 via an overloadinterrupting slide clutch 132, a small bevel gear 134 mounted onto thetransmission shaft 133, a large bevel gear 135 that engages with thesmall bevel gear 134, and a tool holder 137 that is caused to rotatetogether with the large bevel gear 135 in a vertical plane. The hammerdrill 101 can be switched between hammering mode and hammer drill mode.In the hammering mode, the hammer drill 101 performs a hammeringoperation on a workpiece by applying only a striking force to the hammerbit 119 in its axial direction. In the hammer drill mode, the hammerdrill 101 performs a hammer drill operation on a workpiece by applying astriking force in the axial direction and a rotating force in thecircumferential direction to the hammer bit 119.

The striking mechanism 115 includes a striker 143 that is slidablydisposed together with the piston 129 within the bore of the cylinder141, and an impact bolt 145 that is slidably disposed within the toolholder 137 and transmits the kinetic energy of the striker 143 to thehammer bit 119.

As shown in FIGS. 2 and 3, the dynamic vibration reducer 151 isconnected to the body 103 of the hammer drill 101 according to thisembodiment. FIGS. 2 and 3 show the dynamic vibration reducer 151 and thesecond motion converting mechanism 116 that forcibly excites the dynamicvibration reducer 151, in sectional plan view. The dynamic vibrationreducer 151 mainly includes a cylindrical body 152 that is integrallyformed with the body 103 or more specifically the gear housing 107, aweight 153 disposed within the cylindrical body 152, and biasing springs157 disposed on the right and left sides of the weight 153. Each of thebiasing springs 157 is a feature that corresponds to the “elasticelement” according to the present invention. The biasing springs 157exert a spring force on the weight 153 toward each other when the weight153 moves in the longitudinal direction of the cylindrical body 152 (inthe axial direction of the hammer bit 119). Further, an actuationchamber 156 (which is shown in FIG. 2 as right and left actuationchambers 156 a and 156 b) is defined on the both sides of the weight 153within the cylindrical body 152. The actuation chamber 156 communicateswith the outside of the dynamic vibration reducer 151 via an opening 152a formed through the wall of the cylindrical body 152 or via a vent 152b.

The weight 153 has a large-diameter portion 154 and a small-diameterportion 155 which are contiguous to each other. The dimensions of theweight 153 can be appropriately adjusted by selecting, for example, thecontour and the longitudinal length of the large- and small-diameterportions, so that the weight 153 can be made compact in size. Further,the weight 153 is elongated in the direction of its movement and theouter periphery of the small-diameter portion 155 is in close contactwith the inner periphery of the biasing spring 157. Therefore, theweight 153 can be stabilized in movement in the axial direction of thehammer bit 119.

Further, the dynamic vibration reducers 151 are disposed on the bothsides of the axis of the hammer bit 119. In this embodiment, thecylindrical body 152 of each of the dynamic vibration reducers 151 isintegrally formed with the body 103 (the gear housing 107), but it maybe designed to be detachable from the body 103.

The second motion converting mechanism 116 is provided as a means foractively driving the dynamic vibration reducer 151 by forcibly excitingthe weight 153. As shown in FIGS. 2 and 3, the second motion convertingmechanism 116 mainly includes an eccentric shaft 161 formed on thedriven gear 123 (see FIG. 1) of the first motion converting mechanism113, a connecting plate 163 that reciprocates in the axial direction ofthe hammer bit 119 by rotation of the eccentric shaft 161, and a slider167 that linearly moves together with the connecting plate 163 andinputs the excitation force to the biasing spring 157. The eccentricshaft 161 and the connecting plate 163 form a crank mechanism which is afeature that corresponds to the “second driving mechanism” according tothis invention.

The eccentric shaft 161 has a circular cross section having its centerdisplaced a predetermined distance from the center of rotation of thedriven gear 123. The connecting plate 163 engages with the eccentricshaft 161 via an elliptical hole 162 and is guided to move linearly by aplurality of guide pins 165 mounted on the gear housing 107. Further,linkage portions 164 are provided on the both sides of the connectingplate 163 in the direction crossing the direction of movement of theconnecting plate 163 and protrude laterally. Each of the linkageportions 164 extends into the cylindrical body 152 through an opening152 a formed through the wall of the cylindrical body 152 of the dynamicvibration reducer 151 and engages with an engagement recess 168 of theslider 167 disposed within the cylindrical body 152. The slider 167supports one end (on the side of the end of the cylindrical body) of oneof the biasing springs 157 and is disposed for sliding movement in thelongitudinal direction of the cylindrical body 152.

Operation of the hammer drill 101 constructed as described above willnow be explained. When the driving motor 111 (shown in FIG. 1) isdriven, the rotating output of the driving motor 111 causes the drivinggear 121 to rotate in the horizontal plane. When the driving gear 121rotate, the crank plate 125 revolves in the horizontal plane via thedriven gear 123 that engages with the driving gear 121. Then, the crankarm 127 moves in the axial direction of the hammer bit 119 whileswinging in the horizontal plane, and the piston 129 mounted on the endof the crank arm 127 slidingly reciprocates within the cylinder 141. Thestriker 143 reciprocates within the cylinder 141 at a speed higher thanthe reciprocating movement of the piston 129 by the action of the airspring function within the cylinder 141 as a result of the slidingmovement of the piston 129. At this time, the striker 143 collides withthe impact bolt 145 and transmits the kinetic energy caused by thecollision to the hammer bit 119. When the hammer drill 101 is driven inhammer drill mode, the driving gear 121 is caused to rotate by therotating output of the driving motor 111, and the transmission gear 131that engages with the driving gear 121 is caused to rotate together withthe transmission shaft 133 and the small bevel gear 134 in a horizontalplane. The large bevel gear 135 that engages with the small bevel gear134 is then caused to rotate in a vertical plane, which in turn causesthe tool holder 137 and the hammer bit 119 held by the tool holder 137to rotate together with the large bevel gear 135. Thus, in the hammerdrill mode, the hammer bit 119 performs a hammering movement in theaxial direction and a drilling movement in the circumferentialdirection, so that the hammer drill operation is performed on theworkpiece.

When the hammer drill 101 is driven in hammer mode, a midpoint in thepower transmission system, or more specifically, a clutch mechanism 136disposed between the large bevel gear 135 and the tool holder 137 is cutoff. In other words, in the hammer mode, the hammer bit 119 onlyperforms a hammering movement in the axial direction, so that thehammering operation is performed on the workpiece.

As described above, the dynamic vibration reducer 151 mounted in thebody 103 serves to reduce impulsive and cyclic vibration caused when thehammer bit 119 is driven. Specifically, the weight 153 and the biasingsprings 157 serve as vibration reducing elements in the dynamicvibration reducer 151 and cooperate to passively reduce vibration of thebody 103 of the hammer drill 101 on which a predetermined outside force(vibration) is exerted.

In this embodiment, when the hammer drill 101 is driven and theeccentric shaft 161 of the driven gear 123 rotates in the horizontalplane, the connecting plate 163 engaged with the eccentric shaft 161reciprocates in the axial direction of the hammer bit 119. When theconnecting plate 163 moves in one direction (toward the hammer bit 119in this embodiment), the connecting plate 163 moves the slider 167 andpresses the biasing spring 157, which in turn moves the weight 153 inthe direction of pressing the biasing spring 157. Thus, the weight 153of the dynamic vibration reducer 151 is actively driven.

In this embodiment, a spring receiving member in the form of the slider167 is driven via the crank mechanism formed by the eccentric shaft 161and the connecting plate 163 in order to forcibly excite the weight 153.Therefore, the timing for driving the weight 153 or the phase of thecrank can be adjusted such that, in actual design, the weight 153 of thedynamic vibration reducer 151 reciprocates in a direction opposite tothe striker 143 when the striker 143 collides with the impact bolt 145and reciprocates in such a manner as to apply its impact force to thehammer bit 119. Therefore, the vibration of the hammer drill 101 can beeffectively alleviated or reduced.

Further, in this embodiment, after the piston 129 starts to move towardthe striker 143, the striker 143 actually starts to move linearly towardthe impact bolt 145 with a slight time delay due to the compression timerequired for actuation of the air spring, the inertial force of thestriker 143 or other similar factors. Therefore, preferably, the time atwhich the slider 167 presses the biasing spring 157 or the time at whichan excitation force is inputted to the weight 153 is appropriately setallowing for such time delay.

Further, in this embodiment, the actuation chamber 156 is normally incommunication with the outside so that air can freely flow in and out.Therefore, the reciprocating movement of the weight 153 in a directionopposite to the striker 143 is not prevented by the air flow.

According to this embodiment, in a passive vibration reducing mechanismin the form of the dynamic vibration reducer 151, the weight 153 isactively driven by the second motion converting mechanism 116 toreciprocate in a direction opposite to the reciprocating direction ofthe striker 143. Therefore, the dynamic vibration reducer 151 can besteadily operated regardless of the magnitude of vibration which actsupon the hammer drill 101. In other words, the weight 153 of the dynamicvibration reducer 151 can be used like a counter weight that is activelydriven by a motion converting mechanism.

In particular, with the construction in which the biasing spring 157 forapplying a biasing force to the weight 153 is mechanically forciblyexcited via the second motion converting mechanism 116, the adjustmentof the timing for driving the weight 153 or the phase adjustment can befreely made. Therefore, the weight 153 is caused to reciprocate in thedirection opposite to the impact force at the time when the impact forceis generated during hammering operation or hammer drill operation of thehammer bit 119, so that the vibration reducing function of the weight153 can be performed in an optimum manner.

When forcible excitation is caused by the slider 167 that linearlydrives the biasing spring 157 for applying a biasing force to the weight153, the amount of travel of the weight 153 becomes very large withrespect to the eccentricity of the eccentric shaft 161 if the forceexcitation frequency of the slider 167 is in the region of[1/(2π)·(2k/m)^(1/2)](Hz), wherein (k) is the spring constant of thebiasing spring 157 and (m) is the mass of the weight 153. By utilizingthis property, the phase of the eccentric shaft 161 is adjusted suchthat the weight 153 moves in a direction opposite to the striker 143.Further, the weight 153 can be moved a greater distance with a smallereccentricity by adjusting the spring constant of the biasing spring 157,the mass of the weight 153 and the eccentricity of the eccentric shaft161. Thus, optimum vibration reduction can be realized. Further, theamount of travel of the slider 167 that is linearly driven to input theexcitation force to the biasing spring 157 can be reduced. Therefore,the installation space for the second motion converting mechanism 116for driving the slider 167 can be saved, so that the hammer drill 101can be effectively reduced in size.

If the dynamic vibration reducer 151 is disposed on one side of the axisof the hammer bit 119, moment will be generated around a vertical axisperpendicular to the axis of the hammer bit 119 when the weight 153 ofthe dynamic vibration reducer 151 is driven. According to thisembodiment, the dynamic vibration reducers 151 are disposed in the samehorizontal plane on the both sides of the axis of the hammer bit 119.Therefore, moments are generated on the both sides around a verticalaxis perpendicular to the axis of the hammer bit 119 by movement of theweight 153 and act upon each other in such a manner as to cancel eachother out. As a result, undesired generation of moment can be minimizedby provision of the dynamic vibration reducer 151.

(Damping Characteristics of the Dynamic Vibration Reducer)

The cylindrical body 152 of the dynamic vibration reducer 151 isconstructed such that the vent 152 b controls the outflow of airpressurized in the second actuation chamber 156 b during reciprocatingmovement of the weight 153, so that a damping force is forcibly appliedto the weight 153. A “forcible excitation model with a damping element”according to the construction of the dynamic vibration reducer 151 ofthis embodiment will now be explained with reference to FIG. 4.

The model shown in FIG. 4 diagrammatically illustrates the constructionof the dynamic vibration reducer 151 of this embodiment and uses adamping element (the vent 152 b), the biasing springs 157, the weight153 and the second motion converting mechanism 116 (the slider 167). Inthis construction, assuming that a excitation force F₀·cos(ωt+Δ) isinputted from the second motion converting mechanism 116 (the slider167) into the biasing spring 157 (on the right side as viewed in thedrawing), the dynamics can be represented by equations (1) to (4) shownin FIG. 4, and the response of the weight 153 can be represented byequations (5) to (7). In particular, from the equation (5) relating tothe behavior of the weight 153, it is understood that, in theory, theamplitude of the weight 153 is multiplied by a factor of p with respectto the excitation force of the second motion converting mechanism 116and the phase difference between the weight 153 and the second motionconverting mechanism 116 is θ.

On the other hand, when the dynamic vibration reducer 151 is actuallyused, for example, variations in the spring constant of the biasingspring 157, an error in the mass of the weight 153, variations inoperating frequency during operation of the hammer drill 101, etc. mayoccur. In such a case, even if the excitation frequency of the secondmotion converting mechanism 116 is adjusted in response to the operatingfrequency during operation of the hammer drill 101, the amplitude andthe phase difference will vary due to the variations as described above,so that there is a limit in actually ensuring reliable vibrationreducing performance. Therefore, in this invention, both the amplitudeand the phase difference can be stabilized in response to the wideexcitation frequency region, particularly by adjusting the dampingcharacteristics of the weight 153, so that the behavior of the dynamicvibration reducer 151 is stabilized.

Specific steps for obtaining the damping characteristics of the weight153 which are effective in stabilizing the amplitude and the phasedifference in response to the wide excitation frequency region will nowbe explained.

(First Step)

In the first step, the relationship between the coefficient ρ (−) of theamplitude of the weight 153 and the excitation frequency f (Hz) of thesecond motion converting mechanism 116 and the relationship between thephase difference θ (°) between the weight 153 and the vibration inputand the excitation frequency f (Hz) are derived based on equations (1)to (7) in FIG. 4. Further, in the first step, the coefficient p and thephase difference θ are considered by varying only the dampingcoefficient c, in order to obtain a desired damping coefficient c forstabilizing the coefficient ρ and the phase difference θ.

A specific example of the above-described first step will now beexplained with reference to FIGS. 5 and 6. FIG. 5 is a graph of an“embodiment” showing the relationship between the coefficient ρ (−) ofthe amplitude of the weight 153 and the excitation frequency f (Hz) andthe relationship between the phase difference θ (°) between the weight153 and the vibration input and the excitation frequency f (Hz), andFIG. 6 is a graph of a “comparative example” with respect to the“embodiment” shown in FIG. 5.

In the “comparative example” shown in FIG. 6, any region in which both ρand θ are stabilized with respect to the excitation frequency f, i.e. inwhich both of the graphs of ρ and θ become horizontal at the same timein the drawing is hardly found. Therefore, if any variations occur inmanufacturing or in use, actual vibration reducing performance will bedifferent from theoretical setting. As for the specific setting, in thiscase, the mass m of the weight 153 is taken as 64 (g), the springconstant k of the two biasing springs as 7.5 (N/mm), and the dampingcoefficient c as 0.1 (N/m).

In this embodiment, as shown in the “embodiment” of FIG. 5, it isdesigned to create a region in which both ρ and θ are stabilized withrespect to the excitation frequency f by suitably setting a coefficientthat defines the damping characteristics of the weight 153, i.e. thedamping coefficient c in the equation (1) shown in FIG. 4. Specifically,the same values are used for the mass m of the weight 153 and the springconstant k of the two biasing springs as in the “comparative example”shown in FIG. 6, and only the damping coefficient c is changed from 0.1(N/m) to 1 (N/m). In this manner, frequency bands A, B are obtained inwhich both ρ and θ are stabilized with respect to the excitationfrequency f. In this case, when predetermined frequency regions of theexcitation frequency f in the form of the frequency bands A, B cover theactual operating frequency region which is set allowing for variationsin manufacturing or in use of the hammer drill 101, the vibrationreduction by the dynamic vibration reducer 151 is rendered effective.

Further, in this embodiment, ρ and θ can be determined as being stableif the coefficient ρ of the amplitude varies within a specified range(for example, Δρ(A) or Δρ(B) in FIG. 5) with respect to thepredetermined change of the excitation frequency f. In this embodiment,in the frequency bands A, B shown in FIG. 5, the coefficient ρ of theamplitude varies within a specified coefficient range (for example,Δρ(A) or Δρ(B) in FIG. 5) and the phase difference θ between the weight153 and the vibration input varies within a specified range (forexample, Δθ(A) or Δθ(B) in FIG. 5) with respect to the predeterminedchange of the excitation frequency f. Therefore, ρ and θ are determinedas being stable. In this case, the specified ranges Δρ(A), Δρ(B) of thecoefficient ρ of the amplitude of the weight 153 and the specifiedranges Δθ(A), Δθ(B) of the phase difference θ between the weight 153 andthe vibration input can be appropriately set as required, for example,to the specifications of the hammer drill 101. The frequency bands A, Bcorrespond to the “predetermined frequency region”, the amplitude ranges((ρ·F₀) in the equation (5)) corresponding to the specified coefficientranges Δρ(A), Δρ(B) correspond to the “specified amplitude range”, andthe specified phase difference ranges Δθ(A), Δθ(B) correspond to the“specified phase difference range”, according to this invention.Further, preferably, the frequency bands A, B are set allowing forvariations (typically within the range of about 5%) in manufacturing orin use, so as to have a width long enough to cover the range of thevariations. Thus, in the first step, a desired damping coefficient cwhich can stabilize ρ and θ is obtained.

(Second Step)

In the second step, the hammer drill 101 is actually designed inresponse to the damping coefficient c determined in the first step.Specifically, the diameter of the vent 152 b of the cylindrical body 152of the dynamic vibration reducer 151 or the amount of air flow throughthe vent 152 b per unit time is set such that the desired dampingcoefficient c determined in the first step is obtained. The diameter ofthe vent 152 b is typically about 1.0 (mm).

In this embodiment, the design conditions obtained in the first andsecond steps are reflected in the configuration of the vent 152 b of thehammer drill 101. According to the hammer drill 101 of this embodiment,both ρ and θ can be stabilized in response to the wide excitationfrequency range, and the behavior of the dynamic vibration reducer 151can be stabilized even if variations in manufacturing or in use occur.Thus, a reliable vibration reducing performance can be ensured in thehammer drill 101. Further, in this embodiment, the diameter of the vent152 b is set in response to the desired damping coefficient c, which iseffective in simplifying the construction and the design step of thehammer drill 101.

Second Embodiment

A second embodiment of the present invention will now be described withreference to FIGS. 7 to 12. FIG. 7 is a sectional side viewschematically showing an entire electric hammer drill 201 according to asecond embodiment. As shown in FIG. 7, the hammer drill 201 of thisembodiment mainly includes a body 203 and a hammer bit 219 detachablycoupled to the tip end region of the body 203 via a tool holder 237. Thehammer bit 219 is a feature that corresponds to the “tool bit” accordingto the present invention.

The body 203 includes a motor housing 205, a gear housing 207, a barrelsection 217 and a handgrip 209. The motor housing 205 houses a drivingmotor 211 and the gear housing 207 houses a first motion convertingmechanism 213, a power transmitting mechanism 214 and a second motionconverting mechanism 216 (see FIGS. 8 to 12). The barrel section 217houses a striking element 215. The rotating output of the driving motor211 is appropriately converted into linear motion via the first motionconverting mechanism 213 and transmitted to the striking element 215.Then, an impact force is generated in the axial direction of the hammerbit 219 via the striking element 215. Further, the speed of the rotatingoutput of the driving motor 211 is appropriately reduced by the powertransmitting mechanism 214 and then transmitted to the hammer bit 219.As a result, the hammer bit 219 is caused to rotate in thecircumferential direction. Further, the rotating output of the drivingmotor 211 is appropriately converted into linear motion via the secondmotion converting mechanism 213 and inputted into a dynamic vibrationreducer 251, which will be described below, as a driving force forforcibly exciting the dynamic vibration reducer 251. The first andsecond motion converting mechanisms 213 and 216 are features thatcorrespond to the “first operating mechanism” and the “second operatingmechanism”, respectively, according to this invention.

The first motion converting mechanism 213 includes a driving gear 221that is rotated in a vertical plane by the driving motor 211, a drivengear 123 that engages with the driving gear 221, a rotating element 227that rotates together with the driven gear 223 via a driven shaft 225, aswinging ring 229 that is caused to swing in the axial direction of thehammer bit 219 by rotation of the rotating element 227, and a cylinder241 that is caused to reciprocate by swinging movement of the swingingring 229. The driven shaft 225 is disposed parallel (horizontally) tothe axial direction of the hammer bit 219. The outer surface of therotating element 227 fitted onto the driven shaft 225 is inclined at apredetermined angle with respect to the axis of the driven shaft 225.The swinging ring 229 is fitted on the inclined outer surface of therotating element 227 via a bearing 226 such that it can rotate withrespect to the rotating element 227. The swinging ring 229 is caused toswing in the axial direction of the hammer bit 219 by rotation of therotating element 227. Further, the swinging ring 229 has a swinging rod228 extending upward (in the radial direction) from the swinging ring229. The swinging rod 228 is loosely fitted in an engaging member 224that is formed in the rear end portion of the cylinder 241. The rotatingelement 227, the swinging ring 229 and the cylinder 241 form a swingingmechanism, which is a feature that corresponds to the “first drivingmechanism” according to this invention.

The power transmitting mechanism 214 includes a first transmission gear231 that is caused to rotate in a vertical plane by the driving motor211 via the driving gear 221 and the rotating shaft 225, a secondtransmission gear 233 that engages with the first transmission gear 231,a sleeve 235 that is caused to rotate together with the secondtransmission gear 233, and a tool holder 237 that is caused to rotatetogether with the sleeve 235 in a vertical plane. The hammer drill 201of the second embodiment is constructed to perform a hammer drilloperation on a workpiece by applying a striking force to the hammer bit219 in the axial direction and a rotating force in the circumferentialdirection.

The striking mechanism 215 includes a striker 243 that is slidablydisposed within the bore of the cylinder 241, and an impact bolt 245that is slidably disposed within the tool holder 237 and is adapted totransmit the kinetic energy of the striker 243 to the drill bit 219. Thestriker 243 is a feature that corresponds to the “striker” according tothis invention.

FIGS. 8 to 12 show a pair of dynamic vibration reducers 251 and thesecond motion converting mechanism 216 that forcibly excites the dynamicvibration reducers 151. The dynamic vibration reducers 251 are disposedon the both sides of the axis of the hammer bit 119. As shown in FIGS. 8to 10, each of the dynamic vibration reducers 251 mainly includes acylindrical body 252 that is integrally formed with the body 203 or morespecifically the gear housing 207, a weight 253 disposed within thecylindrical body 252, and biasing springs 257 disposed on the both sidesof the weight 253. Each of the biasing springs 257 is a feature thatcorresponds to the “elastic element” according to the present invention.The biasing springs 257 exert a spring force on the weight 253 towardeach other when the weight 253 moves in the longitudinal direction ofthe cylindrical body 252 (in the axial direction of the hammer bit 219).

In this embodiment, the cylindrical body 252 of each of the dynamicvibration reducers 251 is integrally formed with the body 203 (the gearhousing 207), but it may be designed to be detachable from the body 203.

The second motion converting mechanism 216 is provided as a means forinputting excitation force in order to actively drive and forciblyexcite the weight 253 of the dynamic vibration reducer 251. The secondmotion converting mechanism 216 mainly includes a swinging rod 228 forthe swinging ring 229 of the first motion converting mechanism 213, aswinging member 261 that swings together with the swinging rod 228, anoperating piece 263 mounted on the swinging member 261, and a slider 267the is caused to linearly move by the operating piece 263 andmechanically excites one of the biasing springs 257 of the dynamicvibration reducer 251. The swinging ring 229, the swinging member 261and the operating piece 263 form a swinging mechanism, which is afeature that corresponds to the “second driving mechanism” according tothis invention.

As shown in FIG. 12, the swinging member 261 is generally semicircularand disposed astride the upper side of the swinging ring 229. Further, acentral portion 261 b of the swinging member 261 in the circumferentialdirection is fitted onto the swinging rod 228 for relative rotation onthe axis of the swinging rod 228. Further, circular stems 261 a areformed on the both ends of the swinging member 261 and supported by aholder 265 such that it can rotate on a horizontal axis perpendicular tothe axis of the driven shaft 225. Therefore, when the swinging ring 229swings, the swinging member 261 swings on the stems 261 a in the axialdirection of the hammer bit 219.

The slider 267 of the dynamic vibration reducer 251 is fitted into thecylindrical body 252 such that it can slide in the longitudinaldirection of the cylindrical body 252 (in the axial direction of thehammer bit 219). The slider 257 supports one end of one of the biasingsprings 257. The both ends of the swinging member 261 are opposed to theassociated sliders 267, and the operating piece 263 is provided on eachof the ends. The end of the operating piece 263 is in contact with theback of the spring support surface of the slider 267 and moves theslider 267 in a direction of pressing the biasing spring 257.

An actuation chamber 256 is defined on the both sides of the weight 253within the cylindrical body 252. The actuation chamber 256 communicateswith the outside of the dynamic vibration reducer 251 via a vent 252 aformed through the wall of the cylindrical body 252 or via a vent 267 aformed through the slider 267. Thus, the actuation chamber 156 isnormally in communication with the outside so that air can freely flowin and out. Therefore, the reciprocating movement of the weight 253 in adirection opposite to the striker 243 is not prevented by the air flow.

Further, the slider 267 has a cylindrical shape elongated in thedirection of movement and having a closed end in the direction ofmovement. Therefore, the slider 267 can have a wider sliding contactarea without increasing the longitudinal length of the cylindrical body252. Thus, the movement of the slider 267 in the longitudinal directioncan be stabilized.

Operation of the hammer drill 201 of the second embodiment constructedas described above will now be explained. When the driving motor 211(shown in FIG. 7) is driven, the rotating output of the driving motor211 causes the driving gear 221 to rotate in a vertical plane. When thedriving gear 221 rotate, the rotating element 227 is caused to rotate ina vertical plane via the driven gear 223 that engages with the drivinggear 221 and the driven shaft 225. The swinging ring 229 and theswinging rod 228 then swing in the axial direction of the hammer bit219. Then the cylinder 241 is caused to linearly slide by the swingingmovement of the swinging rod 228. By the action of the air springfunction within the cylinder 241 as a result of this sliding movement ofthe cylinder 241, the striker 243 reciprocates within the cylinder 241at a speed higher than the reciprocating movement of the cylinder 241.At this time, the striker 243 collides with the impact bolt 245 andtransmits the kinetic energy caused by the collision to the hammer bit219.

When the first transmission gear 231 is caused to rotate together withthe driven shaft 225, the sleeve 235 is caused to rotate in a verticalplane via the second transmission gear 233 that engages with the firsttransmission gear 231, which in turn causes the tool holder 237 and thehammer bit 219 held by the tool holder 237 to rotate together with thesleeve 235. Thus, the hammer bit 219 performs a hammering movement inthe axial direction and a drilling movement in the circumferentialdirection, so that the hammer drill operation is performed on theworkpiece.

As described above, the dynamic vibration reducer 251 mounted in thebody 203 serves to reduce impulsive and cyclic vibration caused when thehammer bit 219 is driven. When the hammer drill 201 is driven and theswinging ring 229 swings, the swinging member 261 swings in the axialdirection of the hammer bit 219. Then the operating piece 263 on theswinging member 261 vertically swings. When the operating piece 263swings in one direction (downward in this embodiment), the operatingpiece 263 linearly moves the slider 267 of the dynamic vibration reducer251 and presses the biasing spring 257, which in turn moves the weight253 in the direction of pressing the biasing spring 257. Specifically,the weight 253 can be actively driven and forcibly excited. Therefore,like the first embodiment, the dynamic vibration reducer 251 can besteadily operated regardless of the magnitude of vibration which actsupon the hammer drill 201.

Further, with the construction in which the biasing spring 257 ismechanically forcibly excited by the second motion converting mechanism216, like in the first embodiment, the adjustment of the timing fordriving the weight 253 or the phase adjustment can be freely made. Thus,the vibration reducing function of the weight 253 can be performed in anoptimum manner.

The construction and the designing technique of the damping mechanismusing the vent 152 b in the first embodiment can also be applied as-isin the second embodiment.

In the first and second embodiments, the hammer drills 101, 201 aredescribed as a representative example of the power tool in the presentinvention. However, the present invention is not limited to the hammerdrills 101, 201, but may be applied to hammers and also to any powertool which performs an operation on a workpiece by linearly moving atool bit, suitably including a jigsaw and a reciprocating saw whichperform a cutting operation on a workpiece by reciprocating a saw blade.

(Modification to the Elastic Element)

In the first and second embodiments, the biasing springs 157, 257 aredisposed on the both sides of the weights 153, 253. In this respect, asshown in FIGS. 13 and 14, biasing springs 157 a, 257 a may be disposedin actuation chambers 156 a, 256 a formed on the right side of theweights 153, 253, without any spring in actuation chambers 156 b, 256 bformed on the left side. In this case, preferably, provision is made forthe weights 153, 253 to come into surface contact with the inner surfaceof the cylindrical bodies 152, 252, so that the weights 153, 253 canslide with stability. According to this modification, the dynamicvibration reducers 151, 251 can be further simplified in structure.

Description of Numerals

-   101 hammer drill (power tool)-   103 body-   105 motor housing-   107 gear housing-   109 handgrip-   111 driving motor-   113 first motion converting mechanism (first operating mechanism)-   114 power transmitting mechanism-   115 striking element-   116 second motion converting mechanism (second operating mechanism)-   117 barrel section-   119 hammer bit (tool bit)-   121 driving gear-   123 driven gear-   125 crank plate-   126 eccentric shaft (first driving mechanism)-   127 crank arm (first driving mechanism)-   128 connecting shaft-   129 piston (first driving mechanism)-   131 transmission gear-   132 slide clutch-   133 transmission shaft-   134 small bevel gear-   135 large bevel gear-   136 clutch mechanism-   137 tool holder-   141 cylinder-   143 striker-   145 impact bolt-   151 dynamic vibration reducer-   152 cylindrical body-   153 weight-   154 large-diameter portion-   155 small-diameter portion-   156 actuation chamber-   157 biasing spring (elastic element)-   161 eccentric shaft (second driving mechanism)-   162 elliptical hole-   163 connecting plate (second driving mechanism)-   164 linkage portion-   165 guide pin-   167 spring receiving member-   168 engagement recess

1. A power tool comprising: a tool bit, a first operating mechanism thatlinearly drives the tool bit and thereby causes the tool bit to performa predetermined operation, a dynamic vibration reducer that reducesvibration in the operation of the tool bit via a weight thatreciprocates under the action of a biasing force of an elastic element,and a second operating mechanism that mechanically excites the elasticelement to thereby forcibly drive the weight.
 2. The power tool asdefined in claim 1, wherein the elastic element is defined as aplurality of elastic elements that connect the weight to a body of thepower tool, and the second operating mechanism is designed tomechanically excite at least one of the elastic elements.
 3. The powertool as defined in claim 1, further comprising a driving motor, wherein:the tool bit is designed as a hammer bit that performs the operation byapplying a linear impact force to the workpiece, the first operatingmechanism includes a first driving mechanism that converts a rotatingoutput of the driving motor into a linear motion in the axial directionof the tool bit, and a striker that is caused to reciprocate by thefirst driving mechanism and thereby drives the tool bit, and the secondoperating mechanism includes a second driving mechanism that converts arotating output of the driving motor into a linear motion in the axialdirection of the tool bit, and a slider that is caused to reciprocate bythe second driving mechanism and thereby excites the elastic element. 4.The power tool as defined in claim 1, further comprising a drivingmotor, wherein: the tool bit is designed as a saw blade that performs acutting operation on a workpiece by reciprocating movement, the firstoperating mechanism includes a first driving mechanism that converts arotating output of the driving motor into a linear motion in the axialdirection of the tool bit, and a slider that cooperates with the sawblade and is caused to reciprocate by the first driving mechanism,thereby reciprocating the saw blade, and the second operating mechanismincludes a second driving mechanism that converts a rotating output ofthe driving motor into a linear motion, and a slider that is caused toreciprocate by the second driving mechanism, thereby exciting theelastic element.
 5. The power tool as defined in claim 1, wherein thedynamic vibration reducers are disposed on the both sides of an axis ofthe tool bit.
 6. The power tool as defined in claim 1, wherein thedynamic vibration reducer has such damping characteristics that, whenthe second operating mechanism excites the elastic element, theamplitude of the weight varies within a specified amplitude range in apredetermined frequency region of excitation frequencies and that thephase difference between the weight and the second operating mechanismvaries within a specified phase difference range in the predeterminedfrequency region, so that the behavior of the dynamic vibration reduceris stabilized.
 7. The power tool as defined in claim 6, wherein thedynamic vibration reducer has a housing in which the weight is slidablydisposed and a vent that is formed through the housing and providescommunication between an inside region and an outside region of thehousing, thereby allowing air flow between the regions, the amount ofair flow through the vent per unit time being set in response to thedamping characteristics.
 8. A power tool comprising: a tool bit, a firstoperating mechanism that drives the tool bit to reciprocate and therebycauses the tool bit to perform a predetermined operation, a dynamicvibration reducer that includes an elastic element and a weight that canreciprocate under the action of a biasing force of the elastic element,and a second operating mechanism that forcibly excites and drives theweight, wherein the dynamic vibration reducer has such dampingcharacteristics that the amplitude of the weight varies within aspecified amplitude range in a predetermined frequency region ofexcitation frequencies of excitation by the second operating mechanismand that the phase difference between the weight and the secondoperating mechanism varies within a specified phase difference range inthe predetermined frequency region, so that the behavior of the dynamicvibration reducer is stabilized.
 9. The power tool as defined in claim8, wherein the dynamic vibration reducer has a housing in which theweight is slidably disposed and a vent that is formed through thehousing and provides communication between an inside region and anoutside region of the housing, thereby allowing air flow between theregions, the amount of air flow through the vent per unit time being setin response to the damping characteristics.
 10. The power tool asdefined in claim 1, wherein the excitation frequency is substantiallyset at 1/(2π)·(2k/m)^(1/2) Hz, wherein k is the elastic constant of thebiasing spring and m is the mass of the weight in the dynamic vibrationreducer, such that linear momentum of the weight is increased.
 11. Thepower tool as defined in claim 1, further comprising a driving motor,wherein the second operating mechanism includes a crank mechanism thatconverts a rotating output of the driving motor into a linear motion inthe axial direction of the tool bit.
 12. The power tool as defined inclaim 1, further comprising a driving motor, wherein the first operatingmechanism includes a swinging mechanism that converts a rotating outputof the driving motor into a linear motion in the axial direction of thetool bit, and the second operating mechanism includes an operating piecethat is connected to the swinging mechanism and the elastic element inorder to mechanically excite the elastic element by components ofmovement in the axial direction of the tool bit in the swinging movementof the swinging mechanism.
 13. The power tool as defined in claim 12,wherein the operating piece functions as a cam element that transmitsthe components of movement in the axial direction of the tool bit in theswinging movement of the swinging mechanism, to the elastic element.